The present invention relates to hydrodynamic fluid film bearings. In such bearings, a rotating object such as a shaft is supported by a stationary bearing pad via a pressurized fluid such as oil, air or water. Hydrodynamic bearings take advantage of the fact that when the rotating object moves, it does not slide along the top of the fluid. Instead, the fluid in contact with the rotating object adheres tightly to the rotating object, and motion is accompanied by slip or shear between the fluid particles through the entire height of the fluid film. Thus, if the rotating object and the contacting layer of fluid move at a velocity which is known, the velocity at intermediate heights of the fluid thickness decreases at a known rate until the fluid in contact with the stationary bearing pad adheres to the bearing pad and is motionless. When, by virtue of the load resulting from its support of the rotating object, the bearing pad is tilted at a small angle to the rotating member, the fluid will be drawn into the wedge-shaped opening, and sufficient pressure will be generated in the fluid film to support the load. This fact is utilized in thrust bearings for hydraulic turbines and propeller shafts of ships as well as in the conventional hydrodynamic journal bearing.
Both thrust bearings and radial or journal bearings normally are characterized by shaft supporting pads spaced about an axis. The axis about which the pads are spaced generally corresponds to the longitudinal axis of the shaft to be supported for both thrust and journal bearings. This axis may be termed the major axis.
In an ideal hydrodynamic bearing, the hydrodynamic wedge extends across the entire bearing pad face, the fluid film is just thick enough to support the load, the major axis of the bearing and the axis of the shaft are aligned, leakage of fluid from the ends of the bearing pad surface which are adjacent the leading and trailing edges is minimized, the fluid film is developed as soon as the shaft begins to rotate, and, in the case of thrust bearings, the bearing pads are equally loaded. While an ideal hydrodynamic bearing has yet to be achieved, a bearing which substantially achieves each of these objectives is said to be designed so as to optimize hydrodynamic wedge formation.
The present invention relates to hydrodynamic bearings that are also sometimes known as movable pad bearings and methods of making the same. Generally these bearings are mounted in such a way that they can move to permit the formation of a wedge-shaped film of lubricant between the relatively moving parts. Since excess fluid causes undesirable friction and power losses, the fluid thickness is preferably just enough to support the maximum load. This is true when the formation of the wedge is optimized. Essentially the pad displaces with a pivoting or a swing-type motion about a center located in front of the pad surface, and bearing friction tends to open the wedge. When the formation of the wedge is optimized, the wedge extends across the entire pad face. Moreover, the wedge is formed at the lowest speed possible, ideally as soon as the shaft begins to rotate.
The so-called tilt-pad radial bearing is by far the most commonly-prescribed design for machines requiring maximum rotordynamic stability because of its exceptional stability characteristics. Consequently, it has become the standard by which many other radial bearings are measured when seeking a highly stable bearing design. The tilt-pad bearing's popularity is evidenced by the large number of applications found in industry, both as original equipment, and as aftermarket replacements. Applications range from small high-speed machines such as turbochargers and compressors, to very large equipment such as steam turbines and generators. The high rotordynamic stability comes from the reduction of cross-coupled stiffness that occurs when pads are free to tilt about their individual pivot points. This attenuates the destabilizing tangential oil film forces that can induce catastrophic subsynchronous vibration in machines equipped with conventional fixed-geometry bearings. Since so many machines are susceptible to this type of bearing-induced instability, there is a large demand for quality tilt-pad bearings.
Because of its many moving parts and manufacturing tolerances, the tilt-pad design is also the most complex and difficult to manufacture of all journal bearing designs. The design complexity is evident in the number of highly-machined parts required to make up the bearing. Clearance tolerances are additive in the built-up assembly of shell, pivots, and pads, requiring a high degree of manufacturing accuracy to yield acceptable radial shaft clearances. Pad pivot friction under high radial load can also lead to premature wear, or even fatigue failure, which can enlarge clearances and increase rotordynamic unbalance response. All of these requirements combine to make the tilt-pad bearing one which demands maximum attention to design, manufacturing, and materials.
The need for close tolerances manifests itself in known radial pad type bearings because it has been believed necessary to provide an accurately determined clearance between the bearing and the rotating object supported so as to allow the appropriate movement of the bearing pads to form the hydrodynamic wedge. The requirement of close tolerances is particularly troublesome in the manufacture of gas lubricated bearings. Another problem with gas lubricated bearings is the breakdown of the fluid film at high speeds. These problems have limited the use of gas lubricated hydrodynamic bearings.
Moreover, there is still a need for a hydrodynamic radial bearing which can be used in applications where it is essential that the shaft remain centered. Currently, in applications where the shaft can not be allowed to float within a radial envelope, e.g., mechanical seals rotating element bearings are used. In rotating element type bearings, shaft centering is not a problem because the shaft is in effect maintained in solid contact with the housing. With conventional hydrodynamic bearings, however, the shaft is separated from the housing by a spacing known as the radial envelope and in operation the shaft is supported on a fluid film. Thus, because of the spacing between the shaft and the bearing surface in conventional hydrodynamic bearings, the center of the shaft tends to float or drift during operation. In mechanical seals, for example, this movement of the shaft leads to a problem known as "shaft run out" which defeats the operation of the mechanical seal. Alternatives to the commonly used tilt pad bearings have been proposed.
The focus of these attempts has been to provide simple bearing constructions which emulate the performance of more complex tilt pad bearings. For example, on pages 180-181 of Lubrication: Its Principles and Practice, Michell discusses a multiple pad bearing in which the pads are elastically pivoted on an annular member of which they form integral parts. The design shown is extremely rigid because the circumferential dimension of the neck supporting the pads is at least twice as great as the radial dimensions of the neck.
U.S. Pat. No. 3,107,955 to Trumpler discloses one example of a bearing having beam mounted bearing pads that displaces with a pivoting or swing-type motion about a center located in front of the pad surface. This bearing, like many prior art bearings, is based only on a two dimensional model of pad movement. Consequently, optimum wedge formation is not achieved.
In the Hall patent, U.S. Pat. No. 2,137,487, there is shown a hydrodynamic moveable pad bearing that develops its hydrodynamic wedge by sliding of its pad along spherical surfaces. In many cases the pad sticks and the corresponding wedge cannot be developed. In the Greene Patent, U.S. Pat. No. 3,930,691, the rocking is provided by elastomers that are subject to contamination and deterioration.
U.S. Pat. No. 4,099,799 to Etsion discloses a nonunitary cantilever mounted resilient pad gas bearing. The disclosed bearing employs a pad mounted on a rectangular cantilever beam to produce a lubricating wedge between the pad face and the rotating shaft. Both thrust bearings and radial or journal bearings are disclosed.
There is shown in the Ide patent, U.S. Pat. No. 4,496,251 a pad which moves with these web-like ligaments so that a wedge shaped film of lubricant is formed between the relatively moving parts. The use of three spaced ligaments necessarily limits flexibility and prevents simple tilting action.
U.S. Pat. No. 4,515,486 discloses hydrodynamic thrust and journal bearings comprising a number of bearing pads, each having a face member and a support member that are separated and bonded together by an elastomeric material.
U.S. Pat. No. 4,526,482 discloses hydrodynamic bearings which are primarily intended for process lubricated applications, i.e., the bearing is designed to work in a fluid. The hydrodynamic bearings are formed with a central section of the load carrying surface that is more compliant than the remainder of the bearings such that they will move under load and form a pressure pocket of fluid to carry high loads.
It has also been noted, in Ide U.S. Pat. No. 4,676,668, that bearing pads may be spaced from the support member by at least one leg which provides flexibility in three directions. To provide flexibility in the plane of motion, the legs are angled inward to form a conical shape with the apex of the cone or point of intersection in front of the pad surface. Each leg has a section modulus that is relatively small in the direction of desired motion to permit compensation for misalignment. These teachings are applicable to both journal and thrust bearings. While the disclosure of this patent represents a significant advance in the art, it has some shortcomings. One such shortcoming is the rigidity of the support structure and bearing pad which inhibits deformation of the pad surface. Further, the bearing construction is not unitary.
The last two patents are of particular interest because they demonstrate that despite the inherent and significant differences between thrust and journal bearings, there is some conceptual similarity between hydrodynamic journal bearings and hydrodynamic thrust bearings.
This application relates in part to hydrodynamic thrust bearings. When the hydrodynamic wedge in such bearings is optimized, the load on each of the circumferentially spaced bearings is substantially equal.
Presently, the most widely used hydrodynamic thrust bearing is the so-called Kingsbury shoe-type bearing. Like tilt pad radial bearings, the shoe-type Kingsbury bearing is characterized by a complex structure which includes pivoted shoes, a thrust collar which rotates with the shaft and applies load to the shoes, a base ring for supporting the shoes, a housing or mounting which contains and supports the internal bearing elements, a lubricating system and a cooling system. As a result of this complex structure, Kingsbury shoe-type bearings are typically extraordinarily expensive.
An alternative to the complex Kingsbury shoe-type bearing is the unitary pedestal bearings shown in FIGS. 19-20. This bearing has been employed in, among other things, deep well pumps. This relatively simple structure is typically formed by sand casting or some other crude manufacturing technique because heretofore, the specific dimensions have not been deemed important. As shown in FIGS. 19 and 20, the bearing is structurally characterized by a flat base 36PA having a thick inner circumferential projection 38PA, a plurality of rigid pedestals 34PA extending transversely from the base and a thrust pad 32PA centered on each rigid pedestal.
FIG. 20(A) illustrates schematically the movement of the bearing of FIGS. 19-20 in response to movement of the opposing thrust runner in the direction of arrow L. In FIG. 20(A), the deflected position (greatly exaggerated) is illustrated in solid lines and the non-deflected position is illustrated in phantom. The curve PD in FIG. 20(A) illustrates the pressure distribution across the face of the pad. Under load, the thrust pads move around the rigid pedestals in an umbrella-like fashion as shown in FIG. 20(A). By virtue of this umbrella-like movement, only a partial hydrodynamic wedge is formed. Consequently, there is an uneven distribution of pressure across the face of the pad as illustrated in FIG. 20(A). Thus, the bearing has proportionately less hydrodynamic advantage compared to a bearing in which a hydrodynamic wedge is formed across the entire thrust pad face. Moreover, the rigidity of the pedestals and flat inflexible base prevent the movements necessary to optimize wedge formation. The foregoing may explain why bearings of the type shown in FIGS. 19-20, while far less expensive than Kingsbury bearings, have proved less efficient and capable and consequently less successful than the shoe-type bearings.
The present inventor has also discovered that the center pivot nature of both the bearing shown in FIGS. 19-20 and the Kingsbury shoe-type bearing contributes to bearing inefficiency. It should also be noted that, because of their rigid center pivots, neither the Kingsbury shoe-type bearings nor the bearing shown in FIGS. 19-20 can move with six degrees of freedom to optimize wedge formation. Thus, while, in some instances, the prior art bearings are capable of movement with six degrees of freedom, because the bearings are not modeled based upon or designed for six degrees of freedom, the resulting performance capabilities of these bearings are limited.
Prior art hydrodynamic bearings often suffer from fluid leakage which causes breakdown of the fluid film. In radial bearings, the leakage primarily occurs at the axial ends of the bearing pad surface. In thrust bearings, the leakage primarily occurs at the outer circumferential periphery of the pad surface as a result of centrifugal forces action on the fluid. When wedge formation is optimized, fluid leakage is minimized.
Many of today's modern turbomachines, especially those running at high speeds and low bearing loads, require the superior stability characteristics of tilt-pad journal bearings to prevent rotordynamic instabilities. Until now, the design complexity of tilt-pad bearings has precluded their use in many small, high-volume applications where cost and size are important.
Pad bearings with moving supports are also described in U.S. Pat. Nos. 5,054,938 and 5,066,144, both to Russell Ide.
As can be appreciated from the above discussion, the geometry of the pad and pad support structure and the mechanical properties of the bearing material govern bearing performance. In general, the design and optimization of such bearings for a given set of operating conditions necessitates complex and time consuming analysis. While various methods have been proposed for performing bearing design analysis, various drawbacks exist when using these previously proposed methods with pad bearings with moving supports. One drawback of the prior methods is the failure to take surface movement into account. Another drawback is the inefficiency of the prior art methods, due in part to the failure to use a finite element analysis (FEA) technique.